Lloyds Ship Vibration and Noise Guidance Notes

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Ship Vibration and Noise

Guidance Notes

July 2006

Revision 2.1

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Enquires should be addressed to: Peter Filcek

Technical Manager

Technical Investigations (TID) Marine Consultancy Services Lloyd’s Register EMEA 71 Fenchurch Street London EC4M 4BS United Kingdom Telephone: +(44)207 423 1765 Fax: +(44)207 423 1804 Email:

[email protected]

Lloyd's Register is a trading name of the Lloyd's Register Group of entities. Services are provided by members of the Lloyd's Register Group,

for details see www.lr.org/entities.

The Lloyd's Register Group assumes no responsibility and shall not be liable to any person for any loss, damage or expense caused by reliance on the

information or advice in this document or howsoever provided, unless that person has signed a contract with the relevant Lloyd's Register Group entity for the provision of this information or advice and in that case any responsibility or liability is exclusively on the terms and conditions set out in that contract.

Except as permitted under current legislation no part of this work may be photocopied, stored in a retrieval system, published, performed in public, adapted, broadcast, transmitted, recorded or reproduced in any form or by any means, without prior permission of the copyright owner. Enquiries should be addressed to Lloyd's Register, 71 Fenchurch Street, London EC3M 4BS.

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Ship Vibration and Noise Guidance Notes CONTENTS

Section 1. Introduction

1.1 Scope

These Vibration and Noise Guidance Notes are produced to assist in two main areas of shipboard activities:

• Routine measurement surveys • Investigation of problems.

1.2 Basis

These Guidance Notes have been compiled from Lloyd’s Register’s experience of measuring, interpreting and assessing shipboard vibration and noise.

In addition to Lloyd’s Register’s own criteria, certain existing International or National standards have been used in assessment of vibration and noise measurements. The

proposed assessment criteria embody the experience gained in the interpretation, adaptation and extrapolation of these Standards in relation to shipboard conditions.

Reference has also been made to other published information where applicable. These references and other relevant standards are listed in Section 11.

1.3 Application of criteria

1.3.1 These Guidance Notes define the application of proposed criteria for assessing the severity of shipboard vibration and noise in the following areas:

• Accommodation and workspaces, with regard to habitability • Local structural vibration, with regard to risk of cracking

• Machinery vibration, with regard to risk of damage or accelerated wear • Noise, with regard to loss of hearing and ease of verbal communication • Hull surface pressure, with regard to propeller induced excitation.

1.3.2 Differences between ISO 6954:1984 and ISO 6954:2000 Guidelines for the overall evaluation of vibration in merchant ships are covered.

1.3.3 These Guidance Notes do not cover torsional, axial, or lateral vibration of shafting systems.

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SECTION 2 Ship Vibration and Noise Guidance Notes

Section 2. Measurement

2.1 Vibration

2.1.1 Units

Vibration velocity amplitude (± mm/s) is adopted in these Guidance Notes as the principal parameter for evaluating shipboard vibration. Unless specified otherwise vibration amplitude, ± a, is half the peak to peak value of the vibration, Figure 2.1.

Vibration can also be measured in displacement or acceleration. Displacement tends to emphasise the lower frequencies, whilst acceleration emphasises the higher frequencies. The relationship between the three parameters for a sinusoidal waveform is given in Figure 2.1. The preferred units of measurements are:

• displacement, ± mm • velocity, ± mm/s • acceleration, ± m/s2_.

Acceleration is also commonly expressed as a ratio of the acceleration due to gravity, “g”, where 1 g = 9.81 m/s2_{. The abbreviation “Gal”, for Galileo, is sometimes encountered,} where 1 Gal = 1 cm/s2_{= 10}-2_m/s2_.

Figure 2.1 Vibration relationships for sinusoids.

Root mean square (r.m.s.), average and peak values are related by:

value peak 2 1 value .average 2 2 value r.m.s. = π = . . . .ms r peak factor crest =

For a sinusoidal waveform:

displacement, y = a.sin(2πn.t) velocity, v = dy/dt = (2πn).a.cos(2πn.t) acceleration, f = dv/dt = d2_y/dt2 = -(2πn)2_{.a.sin(2πn.t)} where n = frequency, Hz t = time, seconds. A m pli tud e Time r. m .s . aver ag e p eak t o pe ak = 2a pe ak, ± a

2.1.2 Peak and r.m.s. values

Hull structure, habitability and reciprocating machinery criteria use the peak vibration amplitude.

The root mean square (r.m.s.) value is related to the energy content and is used in measurements of rotating machinery vibration.

Lloyd’s Register 4

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In habitability and comfort measurements to ISO 6954:2000 an overall weighted r.m.s. velocity is used. In this case a frequency dependant weighting function is applied to the measurement signal which is intended to modify the data in a way which represents the human perception of whole-body vibration at discrete frequencies in the range 1 to 80 Hz as detailed in ISO 2631-2.

The units should be clearly identified.

2.1.3 Broadband and narrowband

Vibration amplitudes are often measured using a simple digital or analogue meter giving a single value representing either the peak or r.m.s. amplitude across a range of frequencies (defined by the characteristics of the transducer and meter). This is known as the overall or broadband value. Conversely, a narrowband measurement is one which is limited to a small range of frequencies usually centred on a frequency of interest. The smallest width is determined by the resolution of the analyser.

2.1.4 Conversion of measurements

It is not possible to convert from broadband r.m.s. to peak or vice-versa unless the

individual frequencies and amplitudes (also phase angles for r.m.s. to peak) are known. In these cases a detailed narrowband analysis is required. The relationship between an overall r.m.s. value xoa.rms, the r.m.s. components x1.rms, x2.rms, x3.rms..., and the peak amplitudes x1.pk, x2.pk, x3.pk..., at frequencies 1, 2, 3,..., is as follows:

The measurement unit should be selected with care, especially when comparison is to be made against these Guidance Notes or international standards.

(

K

)

K

+

=

+

⎟⎟

⎠

⎞

⎜⎜

⎝

⎛

+

⎟⎟

⎠

⎞

⎜⎜

⎝

⎛

+

⎟⎟

⎠

⎞

⎜⎜

⎝

⎛

=

+

=

2 . 3 2 . 2 2 . 1 2 . 3 2 . 2 2 . 1 2 . 3 2 . 2 2 1.rms .

2

1

2

pk pk pk pk pk pk rms rms rms oa

x

2.1.5 Crest factor

The crest factor is a number relating the maximum amplitude of the waveform to the r.m.s. value. Shipboard vibration signals commonly have crest factors between 2 and 4 if the propeller is the predominant excitation. For a regular sinusoidal waveform the crest factor is

2

as defined in Figure 2.1. 2.1.6 Transducers and filters

Measurements should be made with an electronic system employing transducers which generate signals proportional to velocity or acceleration. Integrators may be used for conversions of velocity signals to displacement, or acceleration signals to velocity or displacement.

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Transducers should be mounted using permanent magnets, studs, hard glue or beeswax. Mounting surfaces should be clean and free of debris, paint, rust, etc. Handheld probes are not recommended for single measurement applications where good accuracy is required. They may however be used in certain monitoring applications where care is taken to ensure repeatability.

Filters may be used to restrict the frequency range of broadband measurements. They should be used with care to avoid attenuation and phase change to signals.

2.1.7 Standards

The specifications for vibration transducers, filter characteristics, signal conditioning, display and recording equipment and calibration procedures should conform to International standards as listed in Section 11.

2.1.8 Calibration

The measuring system should be calibrated in all vibration units of interest before and after the measurements. The calibration should be traceable to national standards.

The characteristics of the measuring system shall be known from calibration with regard to the following:

• frequency response

• effect of transducer orientation and cable length • temperature and other environmental conditions.

2.1.9 Records

Permanent records of vibration measurements may be in the form of: • Binary/ASCII computer files

• Analogue/digital magnetic tapes

• Plots of vibration spectra from a narrowband frequency analysis • Thermal/ultra-violet oscillograph paper

2.2 Noise

2.2.1 Sound pressure level, Lp

Sound pressure levels (SPL) are measured in decibels (dB) using a logarithmic scale where the sound level Lp is given by:

where p is the measured root mean square sound pressure level in Pascals and po is the reference sound pressure level, 20 x 10-6_{Pa. Note: 1 Pascal = 1 N/m}2_{. The relationship} between dB and the pressure ratio p/po is given in Table 2.1.

A subjective assessment of changes in sound pressure levels is given in Table 2.2. and typical sound pressure levels in dBA (see para. 2.2.3) are given in Table 2.3.

dB p p log 20 = L o 10 p Lloyd’s Register 6

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Table 2.1 Conversion of dB and pressure ratio.

Pressure

Ratio - dB + Pressure Ratio Pressure Ratio - dB + Pressure Ratio

1.000 0 1.000 0.0316 30 31.62 0.891 1 1.122 0.0100 40 100 0.708 3 1.413 0.0032 50 316.2 0.501 6 1.995 0.0010 60 1000 0.316 10 3.162 0.0001 80 104 0.100 20 10.000 10-5 _{100 10}5

Table 2.3 Typical sound pressure levels.

Sound level Description

0 dBA Quietest perceivable sound 20-30 dBA Countryside at night, quiet bedroom 30-40 dBA Whispered conversation

40-50 dBA Domestic living room 50-60 dBA General office

60-70 dBA Face to face conversation

70-80 dBA Town centre traffic, domestic vacuum cleaner 80-90 dBA Train at 100 m., engine room flats

90-100 dBA Heavy engineering works, near cylinder heads on slow speed diesel at full power

100-110 dBA Near pneumatic drill, grinder 125 dBA Pain threshold

Table 2.2 Subjective changes in pressure levels.

Change in level Perceived effect

3 dBA just noticeable

6 dBA noticeable

10 dBA twice as loud

Table 2.4 Octave band frequencies and A-weighting Octave band nominal centre frequency Hz Lower passband frequency Hz Upper passband frequency Hz A-weighting correction dB 31.5 22 45 -39 63 45 89 -26 125 89 178 -16 250 178 355 -9 500 355 708 -3 1,000 708 1,410 0 2,000 1,410 2,820 +1 4,000 2,820 5,620 +1 8,000 5,620 11,200 -1 16,000 11,120 22,400 -7 2.2.2 Frequencies of interest

The normal frequency range of hearing of young adults is approximately 20 to 18,000 Hz. The frequency range of human speech is principally between 350 and 3,500 Hz.

2.2.3 A-weighted sound pressure level, LpA.

Sound pressure level measurements relating to effect of noise on humans should be made in decibels using an A-weighting filter, dBA. The A-weighted frequency filter is used to reproduce the frequency response of the human ear. The A-weighting corrections are given in Table 2.4.

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2.2.4 Equivalent continuous sound level, L Aeq,T

This is the notional A-weighted, continuous, steady sound pressure level that, within a specified time interval T, has the same mean square sound pressure level as a sound whose level varies with time. It is defined by:

( )

⎥⎦

⎢⎣ −

∫

1 2 0 1 2 10 , t T Aeq

p

t

⎥

⎤

⎢

⎡

2

1 log

10 =

2 t A

dt

t

p

L

where: t2 - t1 = time interval, T

pA(t) = instantaneous A-weighted sound pressure level

p0 = reference sound pressure level

For a continuous, unvarying noise, LAeq is numerically the same as LpA. It is therefore possible to use a conventional sound level meter to determine LAeq if the noise levels for the whole period vary by less than 5 dB with a slow meter response.

2.2.5 Measurement technique

Measurements can be made using LAeq (or Leq as necessary) if the sound level meter has the capability. This is often easier and less subjective than using LpA or Lp. LAeq should be used whenever the LpA fluctuates by more than 5 dB, or the sound is cyclic, irregular, or

intermittent.

Readings should made to the nearest decibel. The microphone should be at head height, 1.2-1.6 metres above deck, and pointed towards the dominant noise source if any.

Measurements should normally be taken in the middle of spaces and no closer than 1 metre from bulkheads and major reflecting sources.

If measurements are made using LpA or Lp, the meter should be set to “slow” response except where noise levels approach the overriding limits. A measuring time of at least 5 seconds should be allowed. If the level fluctuates by no more than 5 dB, an average of the maximum and minimum excursions can be made by eye.

2.2.6 Octave band filters

If the maximum noise level for a space is exceeded or if subjectively annoying tones are present, then the unweighted noise level in each of the octave bands should be determined. The levels in each band are used to determine the Noise Rating number, Section 3.4.

The octave band frequencies cover the normal range of human hearing frequencies and are listed in Table 2.4. The bands have a ratio of upper to lower frequencies of 2 and are centred on preferred frequencies given in ISO 266. Octave bands are comparatively coarse and use is sometimes made of third-octave bands.

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2.2.7 Equipment standards

Sound level meters can be either precision or industrial grade. Meters should conform to international standards, for example IEC 651 type or better and IEC 804 type 2 or better for integrating-averaging meters. Precision meters have an accuracy of about ±1 dB; industrial grade meters have an accuracy of about ±3 dB. A factor of 3 dB should be added to

industrial grade meter readings to cater for the reduced accuracy. The use of a precision grade meter is, therefore, recommended where the noise levels are likely to be close to the recommended levels or in cases of dispute.

Octave filter sets should conform to IEC 225 or equivalent.

Microphones should be of the random incidence type and should conform to IEC 179 and IEC 651 (Types 1 and 2) or equivalent.

2.2.8 Calibration

A suitable calibrator, approved by the manufacturer of the particular sound level meter, should be used. Calibrators for precision meters should be accurate to ± 0.3 dB and for industrial meters to ± 0.5 dB.

The sound level meter, octave filter set and calibrator should be returned to the manufacturer or other organisations which provide a calibration traceable to national standards at intervals not exceeding two years.

2.2.9 Wind screens

A microphone windscreen should be used when taking readings on bridge wings, on open deck, or below decks where there is any substantial air movement.

2.2.10 Recordings

A tape recorder having a linear response in the frequency range 20 to 20,000 Hz should be used in cases of dispute or investigation work.

2.2.11 Speech intelligibility

Speech intelligibility over public address and similar systems may be determined using techniques such as the Speech Interference Level (SIL) or Rapid Speech Transmission Index (RASTI). See also Chapter 9.

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2.3 Hull surface pressures

2.3.1 Pressure transducers

The pressure transducers, signal conditioning, and recording system should have a frequency response range which will measure the impulsive pressures arising from cavitation. An upper frequency response of about 5,000 Hz should be adequate for most purposes.

Corrosion resistant pressure transducers should be used. Ideally the sensitive membrane of the transducer should be flush with the outer hull surface in order to avoid unwanted pressure harmonics. The design of available transducers and fitting sometimes make this difficult to achieve in practice.

2.3.2 System calibration

The pressure transducers and complete recording system should be calibrated before and after the measurements.

2.3.3 Pressure components

The hull surface pressure comprises two components. The first is the direct radiated pressure from the propulsor and the second is a self induced pressure resulting from the vibration of the transducer mounted on the hull.

To separate these two components it is necessary to measure the vibration at the pressure transducer locations. This data, via suitable transformation methods, can be used to estimate the self-induced pressure components in terms of amplitude and phase at the transducer location.

2.3.4 Measurement locations

The number of pressure transducers should, ideally, be between five and seven. For a right-handed propeller of diameter D, four pressure transducers should be placed at 0.05D to starboard of the shaft centreline, Figure 2.2. The longitudinal positions should be in the measurement reference plane shown in Figure 2.2, at intervals of 0.15D, starting at 0.10D aft of the propeller tip plane. At the plane 0.05D ahead of the propeller tips additional

transducers may ideally be placed at 0.10D to port and 0.15D and 0.25D to starboard. The mirror image of this pattern should be applied for a left handed propeller.

For ships with significant areas of shell plating aft of the propeller plane, pressure transducers may also be required to be located at distances up to 2D aft of the propeller plane in line with the principal tip vortex activity in the wake peak.

2.3.5 Phase marker

A phase marker or angular position indicator should be fitted to the inboard shafting. It is convenient if this coincides with a particular blade at a known angular position of the propeller.

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2.3.6 Visual observation

Consideration should also be given to fitting viewing ports in the hull to allow boroscopes or video cameras and stroboscopic lighting to be used for observation in cases where severe cavitation occurs.

Figure 2.2 Propeller pressure measurement positions.

0.05D

Measurement reference plane

0.05D 0. 8 D iam et e r 0.10D 0.10D 0.10D Mid-chord 0.10D

Drawn for right-hand propeller, looking from above.

Positions for left-hand propeller are reversed in mirror image about shaft centreline.

Starboard 0.15D Port Shaft centreline 0.15D Lloyd’s Register 11

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Section 3. Analysis

3.1 Vibration

3.1.1 Narrowband analysis

Narrowband frequency analysis is the recommended method of interpreting vibration measurements in relation to the guidance limits or for investigation purposes.

Narrowband frequency analysis should be capable of resolving the frequency of individual components of interest over the relevant frequency range.

3.1.2 Broadband limits

Some guidance values based on broadband measurements of the overall amplitude of composite frequency vibration are suggested in later sections to simplify routine survey procedures.

Broadband measurements should cover the specified range of vibration frequencies for each particular application.

If the acceptability of vibration severity is marginal from the results of broadband

measurements, narrow band frequency analysis should be used. Records of the vibration measurements should be taken in such cases.

3.1.3 Orders of vibration

Vibration frequencies can conveniently be related to a known fundamental excitation frequency by the use of order numbers. In cases where shaft rotational frequency is an appropriate reference, then:

Order numbers are typically integer numbers, 1, 2, 3,...,n. Vibration at four cycles per shaft revolution arising from a four-bladed propeller, for example, can either be described as occurring at shaft fourth order or at blade order. Half orders (½, 1½, ...) may typically occur with 4-stroke diesel engines where the working cycle takes place over two shaft revolutions.

frequency rotational Shaft frequency Vibration number Order = 3.1.4 Analysis methods

Vibration data can be divided into two categories: steady and transient. Some data with minimal transient characteristics can be considered in a quasi-steady sense. The available analysis methods are summarized in Figure 3.1.

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Figure 3.1 Summary of available analysis methods Non-steady

signal

Joint Time Frequency (JTFA) Data dependent Power spectra Fourier analysis (FFT) Short time Fourier transforms (STFT) Gabor spectrograms Adaptive spectrograms Spectrograms (Linear JTFA) Time series Steady signal Daubechies Haar Specific design Distributions (Bilinear JTFA) Cohen's Class Wavelets Wigner-Ville Choi-Williams Cone-shped Other Time scale

3.1.5 Fast Fourier Transform analysis

Analysis of steady and quasi-steady signals is usually carried out using the Fast Fourier Transform (FFT) function found on most vibration analysers.

An FFT analysis transforms consecutive samples (typically 1024, 2048, or 4096) from the time domain to the same number of lines in the frequency domain. The algorithm used to

calculate the FFT is finite and discrete. This has a number of effects.

The first is that aliasing might occur, when because of a limited number of samples, high frequencies might appear as false lower frequencies. This is avoided by lowpass “anti-aliasing” filters. A sampling frequency higher than the maximum frequency of interest, fmax , is used. This is determined by:

2

max

frequency sampling

f =

The frequency resolution is thus dependent on the frequency range used, but most analysers have band- selectable analysis which gives increased resolution or “zoom” at frequencies of interest.

The second effect of a finite sample record is that discontinuities can arise at the ends of the sample, giving rise to false results when the signal is not periodic in the time record. This causes leakage of energy from one resolution line of the FFT into other lines. The amplitude at the start and end of each block of samples is therefore reduced to zero, a technique known as windowing. The Hanning window is a good, general purpose compromise for continuous signals. A transient signal, such as from an impact, is self-windowing and a Uniform (Rectangular) window should be used.

The third effect is known as the picket fence effect. It arises from the discrete sampling of the spectrum in the frequency domain in which the FFT acts like a series of parallel filters. The results are similar to viewing the results through slits in a picket fence.

The shape or response of the filters is determined by the window function used. The amplitude of a frequency which is in the middle of a filter using a Hanning window will be measured accurately. A frequency midway between filters could be attenuated by up to 1.5 dB (18%). The use of a Flat-top window will reduce the possible attenuation to less than 0.1 dB (1%). However, the increase in accuracy comes at the expense of frequency

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Time domain signal

It is desirable to view the “raw” vibration signal (signal amplitude displayed against time) during analysis. Signal characteristics such as beating, transients and irregularities can be readily detected in the time domain. Approximate inspection methods of analysis are recommended as a check on the results derived from analytical techniques.

Averaging

Averaging is not always appropriate and its use may destroy or disguise the original signal characteristics. This includes, for example, signals arising from intermittent defects or averaging of non-steady signals. It may be possible to use non-steady signal techniques to give a better understanding of the signal characteristics. Root mean square (r.m.s.) averaging will give a better estimate of the value of a signal, but it will not improve the signal-to-noise ratio. Linear averaging will improve the signal-to-noise ratio if a trigger signal is available which is synchronous with the periodic part of the signal.

3.1.6 Modulation

Modulation describes a time dependant variation, either random or repetitive, in a vibration signal, Figure 3.2.

• Amplitude modulation arises, for example, in an eccentrically mounted gear where the tooth meshing frequency is constant but the amplitude varies, typically at once per revolution of the gear.

• Frequency modulation occurs, for example, in gears with tooth spacing errors or the passing signal from torsionally vibrating gear teeth.

• Phase modulation occurs, for example, onboard a twin screw ship where the excitation varies in time.

These types of modulation can occur in the same signal, for example in heavy weather where the speed and load of a ship’s main engine may vary simultaneously or in the hunting of a governor of an alternator set. Frequency analysis cannot describe the varying amplitude and frequency. The information is interpreted as steady sine waves which appear as side-bands about the fundamental frequency. The amplitude of the fundamental frequency may be significantly diminished in cases of very heavy modulation.

One side effect of modulation is that the ear may detect frequencies that do not exist or may be outside the normal range of audible frequencies.

Figure 3.2 Vibration modulation

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3.1.7 Rolling element envelope analysis

Incipient defects in rolling element bearings produce sharp pulses which may produce measurable frequencies up to 20 kHz. Identification of the source cannot be carried out using simple FFT analysis and a technique known as envelope analysis is used to extract useful information from the vibration signal as follows. First the time domain signal is band-pass filtered around the region of high frequency energy. This leaves a signal containing bursts of energy at the defect frequency, which is then rectified and low-pass-filtered. An FFT analysis of the resultant signal will allow the defect frequency to be identified. Calculation of the impact rates is given in Figure 3.3.

3.1.8 Calibration

Analysis methods should be verified using known input data. Any software used in calibration, measurement or analysis should be part of an appropriate Quality Assurance system.

Figure 3.3 Calculation of rolling element impact frequencies.

Ball diameter, Contact angle, Pitch diam, β B P

Impact frequencies, f Hz, assuming pure rolling:

angle contact = diameter pitch race = diameter roller or ball = races outer and inner between second per s revolution relative = f rollers or balls of number = n where 2 cos . -1 . . = f defect, roller / Ball cos . + 1 . . 2 = f defect, race Inner cos . -1 . . 2 = f defect, race Outer r β β β β P B P B r f B P P B r f n P B r f n ⎥ ⎥ ⎦ ⎤ ⎢ ⎢ ⎣ ⎡ ⎟ ⎠ ⎞ ⎜ ⎝ ⎛ ⎟ ⎠ ⎞ ⎜ ⎝ ⎛ ⎟ ⎠ ⎞ ⎜ ⎝ ⎛ Lloyd’s Register 15

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3.2 Analysis for ISO 6954:1984

3.2.1 This section details Lloyd’s Register’s interpretation of ISO 6954:1984 Mechanical Vibration

and shock -Guidelines for the overall evaluation of vibration in Merchant Ships (1984). It should be

used in conjunction with Section 6.5.

The analysis is different to that used in ISO 6954:2000 and users should ensure that the appropriate version is being used.

3.2.2 Derivation of Vibration Levels

In order to define the correct procedure for interpretation of FFT analysis, it is useful to consider how the guidance vibration levels of ISO 6954:1984 were originally derived.

The measure of vibration severity used in ISO 6954:1984 is termed the maximum repetitive amplitude (MRA). The MRA is defined as the largest repeating value of the modulated signal of a single frequency obtained during sea trials. The sea trials are to be conducted in accordance with ISO 4867: Code for the measurement and reporting of shipboard vibration data and / or ISO 4868: Code for the measurement and reporting of local vibration data of ship structures

and equipment.

The usual measuring system employed to collect the vibration data used in formulating the original guidance limits of ISO 6954:1984 included a chart recorder that provided a paper trace of the vibration signal. This trace was then manually analysed to determine the MRA for each significant excitation source.

A representation of a typical propeller excited hull vibration is shown in Figure 3.4. It indicates modulation of the vibration amplitude due to variation of the propeller excitation forces. This vibration signal has been filtered to show only propeller blade passing frequency from which the MRA is readily identified. The value indicated is the peak-to-peak value, or twice the MRA. This is the value to be compared with the guidance levels given is ISO 6954:1984.

MRA’s for individual excitation frequencies can be readily evaluated using a sea trials measuring system that includes a chart recorder, a tape recorder and precision filters. However, this is laborious, and an FFT analyser is often used to provide a spectral analysis of the vibration signal. Unfortunately, as a consequence of the averaging process used in an FFT analyser, the modulation effects are suppressed and the MRA is obscured. A direct reading of the MRA as required by ISO 6954:1984 is not possible and it is therefore necessary to apply a correction factor in order to estimate the MRA.

Figure 3.4 Maximum repetitive amplitude

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3.2.3 Use of FFT Analysers

When using an FFT analyser to provide the frequency spectra of a vibration signal there are, on a typical analyser, three analysis methods available, namely:

• r.m.s. average • peak average

• peak hold - not to be used for MRA estimates.

Each of these methods calculates a time-averaged level of vibration over the recorded signal length and, as a consequence of the averaging process, the true MRA cannot be directly calculated. Of these three averaging methods, the analysis procedure that most accurately reflects the true value of the MRA, as defined in ISO 6954:1984, is as follows:

• Calculate the average r.m.s. value of vibration over the frequency range 1 - 100 Hz. The record should be at least two minutes long.

• Multiply the average r.m.s. spectral values by a crest factor of 2.5 for propeller excited vibration to allow direct comparison with typical chart recorder traces or the ISO 6954:1984 guidelines.

The crest factor of 2.5 comprises:

• A factor of 1.414 to convert from r.m.s. to peak values.

• A tentative conversion factor of 1.8, as proposed in ISO 6954:1984, to account for the normal modulation of the vibration signal in sea state 3. The conversion factor is defined as the ratio of MRA to the average amplitude during steady speed trials.

It is stressed that the conversion factor of 1.8 is a tentative value that may vary between ships and/or trial conditions. It is therefore essential that the measuring equipment employed during sea trials provide a recording of the time based vibration signal, either digitally or on magnetic tape, so that if required, the true conversion factor can be calculated. This may be necessary in the event of possible design problems or questions relating to compliance with design specifications. For example, structural vibration due to machinery excitation is unlikely to have the degree of modulation typically associated with propeller excitation and therefore the use of 1.8 as the conversion factor will lead to an over estimate of the MRA.

3.2.4 Recommended analysis procedure

In the examination or the interpretation process of vibration signals, the fundamental step is to identify the excitation sources. For vibrations due to machinery, a factor of 1.414 to convert from r.m.s. values to peak values is normally adequate to assess the vibration severity. For propeller induced vibrations, however, a more cautious approach will be adopted.

In the case where the vibrations caused by propeller excitation are not considered to cause any complaints and that the levels are well within design specifications, the use of a crest factor of 2.5 is acceptable. However, in the event that the vibration levels are marginally within the design specifications and that possible problems or argument may arise, the vibration signals will then be subjected to vigorous examination in order to establish the true conversion factor. This examination process is:

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• to play back the recorded signal from a tape recorder onto an oscillograph measuring device to obtain a paper trace of the vibration signal;

• a ‘filter’ will be used during the above play back process so that only the vibration signals related to the propeller blade passing frequency will be examined. The filter response and the bandwidth should be wide enough to capture the quickest

modulations;

• the recorded signal on the paper trace should be long enough to allow the establishment of MRA and the average amplitude;

• the true crest factor will then be calculated from the MRA and average amplitude.

3.3 Analysis for ISO 6954:2000

3.3.1 This section details Lloyd’s Register’s interpretation of ISO 6954:2000 Mechanical Vibration

and Shock – Guidelines for the overall evaluation of vibration in Merchant Ships (2000). It should

be used in conjunction with Section 6.6.

The analysis is different to that used in ISO 6954: 1984 and users should ensure that the appropriate version is being used.

3.3.2 The vibration level is defined as an overall frequency weighted r.m.s. velocity for all vibrations within a range of 1 to 80 Hz. Acceleration may also be used. The weighting functions, from ISO6954:2000 are reproduced in Table 3.1 and Figure 3.5. The weighting functions are applicable to all directions.

Table 3.1 Weighting functions for ISO 6954:2000.

Band Velocity Acceleration Band Velocity Acceleration

Hz Wv Wa Hz Wv Wa 0.20 0.002 0.063 10.00 0.869 0.494 0.25 0.004 0.099 12.50 0.911 0.412 0.32 0.009 0.156 16.00 0.941 0.337 0.40 0.017 0.243 20.00 0.961 0.274 0.50 0.032 0.369 25.00 0.974 0.220 0.63 0.059 0.530 31.50 0.979 0.176 0.80 0.098 0.701 40.00 0.977 0.140 1.00 0.147 0.833 50.00 0.964 0.112 1.25 0.201 0.907 63.00 0.926 0.083 1.60 0.261 0.934 80.00 0.843 0.060 2.00 0.327 0.932 100.00 0.706 0.040 2.50 0.402 0.910 125.00 0.533 0.024 3.15 0.485 0.872 160.00 0.370 0.013 4.00 0.573 0.818 200.00 0.244 0.007 5.00 0.661 0.750 250.00 0.156 0.004 6.30 0.743 0.669 315.00 0.100 0.002 8.00 0.813 0.582 400.00 0.063 0.001 Lloyd’s Register 18

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Figure 3.5 Weighting functions for ISO 6954:2000.

0.001 0.01 0.1 1 10 0.1 1 10 100 1000 Frequency (Hz) Fr e q u e nc y we ig h ti ng Acceleration weighting, Wa Velocity weighting, Wv 3.3.3 Record length

Analysis of recorded measurements should be made for a period of at least one minute, or two minutes if there are dominant components below 2 Hz. For unsteady vibrations with significant modulation, such as structural response to propeller induced vibration forces, it may be necessary to extend further the measurement and the corresponding analysis periods in order to derive reliable values of overall r.m.s. vibration velocity.

3.3.4 Analysis options

Analysis of the measured data may be undertaken in either the time or frequency domains. The availability of portable FFT signal analysers means that a frequency domain analysis will be preferred by most vibration engineers. The procedure consists of three steps: • spectral analysis (constant or proportional bandwith), followed by

• signal weighting, and finally

• summation of weighted spectral components.

3.3.5 FFT analyser settings

Typical FFT parameters for a 400-line analyser are: • Hanning window

• Frequency range, 0-100 Hz (nearest to 1-80 Hz) • Sampling frequency, 250Hz

• Number of data in each sample: 1024 • Sample length: 4.096 seconds

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SECTION 3 Ship Vibration and Noise Guidance Notes Lloyd’s Register 20 2 / 1 2

⎤

⎡

=

∑

a

3.3.6 Calculation of overall weighted value

The overall weighted value is determined from:

⎥

⎦

⎢

⎣

i wi w Where:

aw = frequency-weighted acceleration or velocity.

awi = weighted acceleration or velocity for the i th one-third octave band

3.3.7 Direction

The highest value in any direction will be used for assessment.

3.4 Noise

3.4.1 Noise Rating Curves

3.4.2 Noise Rating (NR) curves are used in the IMO Code on Noise Levels on board Ships and in statutory codes and were incorporated into ISO 1996. They are a set of empirical curves relating the linear octave band pressure level to the centre frequency of the octave bands each of which is characterised by a Noise Rating number which is numerically equal to the sound pressure level given by the curve at 1000 Hz, Figure 3.6. The subjective effect of the noise is predicted from the NR number. The octave band sound pressure levels are plotted on the NR curves. The NR number is that NR curve to which the highest plotted octave level is anywhere tangent.

3.4.3 Narrowband analysis of noise

The narrowband analysis of noise is useful in locating sources which are dependant upon speed of rotation and give rise to pure tones, for example, fans at blade passing frequency.

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Figure 3.6 Noise rating curves

31.5 63 125 250 500 1000 2000 4000 8000 Octave band centre frequency, Hz

20 30 40 50 60 70 80 90 100 110 120 130 140 O ct a ve ban d pr essur e le vel , dB r e 20 m icr o P a scal 130 125 120 115 110 105 100 95 90 85 80 75 70 65 60 55 50 45 40 35 30 25 20

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3.5 Hull surface pressure data

3.5.1 Cavitation induced signatures can take the form of either discrete harmonics (blade orders) or have a broadband character. The signals are also non-steady and, therefore, the

measurement needs to take account of this. The non-steady character may be a function of any propeller blade to blade geometry differences, temporal and spatial wake fluctuations, and water quality and environmental changes.

3.5.2 Time series data are most commonly reduced to their constituent frequency content using Fast Fourier Transform (FFT) analysis techniques. Such processes generate frequency spectra of pressure amplitudes and phases averaged over the length of the time series records which contain several hundred blade passages. Such a procedure does not distinguish the contributions made by individual blades or account for discrete bursts of activity.

3.5.3 To establish the nature of the pressure signal time histories, in particular beating type effects, the measurements should be viewed as time series to present the relationship between pressure and propeller angular position. The pressure signatures should be considered over both short and long time frames.

3.5.4 To distinguish short time frame events alternative methods of analysis are necessary. Such methods would include Joint Time Frequency Analysis (JTFA) in which segments of the signal are analysed individually and the frequency content is obtained as a function of time. Such procedures can be extended to the separate analysis of the signatures from each blade.

3.5.5 When available, synchronised high-speed video and pressure signals should be presented to assist in connecting cavity and pressure events to identify the source of the higher harmonic content of the signals.

3.5.6 Pressure spectra and 1/3rd octave band analysis should be presented in order to supplement the amplitudes extracted at multiples of the blade passing frequency.

3.5.7 Hull pressures should be given as single amplitude peak values (half the peak-to-peak).

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Section 4. Ship structure vibration

4.1 Hull girder

4.1.1 Resonance

A ship’s hull will respond to exciting forces and moments as a beam freely supported in water. The modes are described by the plane of vibration and the number of nodes (points of no deflection); for example, the two-node vertical, Figure 4.1. Frequencies are low, typically in the range 1-10 Hz. The excitation may come from engine external forces or moments, propeller pressure impulses or, irregularly, by waves. Wave excitation frequencies are typically 1-2 Hz.

Figure 4.1 Hull girder modes of vibration.

2-node vertical or horizontal

3-node vertical or horizontal

Torsional

At resonance, when excitation and hull girder natural frequencies coincide, dynamic magnification may result in large amplitudes of vibration for comparatively low excitation.

Hull girder natural frequencies are dependant on the stiffness and mass of the structure and the virtual added mass of water. They vary with changes in draught, trim, and mass distribution. Resonances above the normal operating speed of the vessel in ballast may move down the speed range and become troublesome in the loaded condition.

Investigation of hull girder vibration is fundamental to identification of the possible causes of high shipboard vibration.

4.1.2 Measurement positions

It is normal to measure hull girder response at the stern or, alternatively, at the sterntube, Figure 4.2. Measurements at the sides will allow torsional hull modes to be identified and simultaneous measurements at the wheelhouse and the main engine (if a slow speed diesel) will provide additional information about superstructure and engine modes.

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Figure 4.2 Measurement positions.

Sterntube

Stern

Bridge

Main engine

4.1.3 Avoidance of resonance in service

Resonances of the ship’s structure are normally lightly damped and finely tuned so that an adjustment of the shaft speed by a few revolutions either way may reduce the response. It may be possible to alter the frequency or magnitude of the excitation source, for example, changing the number of blades on a propeller or changing to a highly skewed or

tip-unloaded propeller blade form. It is generally not practical to change the natural frequencies of hull girder modes by altering the steelwork or to reduce the response by additional damping.

4.1.4 Avoidance of resonance by design

3-D finite element analysis can be used at the design stage to model a ship’s structure. Reliable estimates can be made of natural frequencies which might be excited at service speed. Campbell diagrams are a useful tool here in plotting excitation and response frequencies.

4.1.5 Forced vibration

Forced vibration may occur when propeller pressure impulses or engine external forces and couples are large. Amplitudes increase with transmitted shaft torque and are generally proportional to the square of the speed with a fixed pitch propeller. A reduction in the excitation level is usually necessary if forced vibration levels are excessive.

4.2 Superstructure modes

Superstructure vibration may be caused by vibration of the hull girder or resonant behaviour of the deckhouse, Figure 4.3. A lack of deckhouse stiffness and poor vertical support or continuity will exacerbate vibration levels.

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Figure 4.3 Ship superstructure modes of vibration. Shear deflection

Rigid body - elastic support

Rigid body - hull girder deflection

4.3 Acceptable vibration levels

Excessive vibration of the ship’s structure manifests itself as cracking of the ship’s structure (e.g. bulkheads, tanks), failure of equipment (e.g. radars), and uncomfortable living and working conditions. Guidance values for these as local phenomena are given elsewhere in these Guidance Notes.

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Section 5. Local structural vibration

5.1 General

Guidance values are based on Lloyd’s Register’s full scale experience of vibration induced structural cracking. The principal source of the data concerns the vibration of panels relative to their stiffened boundaries.

Structural fatigue is a complex failure mode and of necessity the guidance values have a broad stippled zone to cater for the wide variations encountered with regard to:

• structural configurations and geometry • vibration mode

• weld details • workmanship • environment.

It should be noted that the natural frequencies of tank panels are significantly influenced by the density and level of liquids within the tank. Damage may therefore only occur during particular cargo or operating conditions.

5.2 Scope

The assessment criteria are restricted to locally magnified vibrations of steel structural elements relative to the boundaries with regard to the risk of cracking. The frequency range considered is 5 to 100 Hz.

5.3 Assessment

The assessment levels are shown in Figure 5.1 and apply to unidirectional narrowband measurements taken at the position of maximum vibration amplitude such as at the centre of panels or tank sides.

It is recommended that the overall vibration amplitudes (peak) for vibrations in the frequency range 5 to 100 Hz are less than each of the following values:

• displacement ± 0.25 mm peak • velocity ± 30 mm/s peak • acceleration ± 20 m/s2_peak.

5.4 Alternative techniques

Structural vibrations may be assessed by dynamic strain measurements in conjunction with other relevant data to define the fatigue environment.

Modal and finite element analyses may give a better understanding of the behaviour of structures. The assumptions implicit in such analyses should be validated by

measurements.

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Figure 5.1 Local structural vibration.

Damage probable 1 1 mm/s 0. 01 m /s² 10 mm/s V e lo ci ty ( p e a k) , ± m m /s Frequency, Hz 10 0.1 m /s² Recommended 100 mm/s 10 m m 0.1 m m 1 m /s ² 100 10 m /s² Disp lacem ent, ± mm Acc eler at ion, ± m/ s² 0.01 m m 100 m /s² 1 m m Lloyd’s Register 27

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Section 6. Accommodation and workspace vibration

6.1 Scope

6.1.1 The scope of this section is the evaluation of hull and superstructure vibration with regard to crew habitability in all normally occupied spaces. The vibration levels given in this section do not apply to passenger cabins or other passenger spaces except in so far as they are working areas.

6.1.2 The evaluation applies to normal service operating conditions with the main propulsion machinery developing at least 85% of the maximum continuous rated power. The evaluation is not intended to apply to manoeuvring conditions or operation in heavy weather.

6.2 Measurements

6.2.1 Measurements should be taken on the deck in three mutually orthogonal directions: vertical, athwartships and longitudinal.

6.2.2 Where normal transducer attachment methods are impractical, due to the presence of deck coverings such as carpets, an inertia block with a mass of at least 1kg with spiked supporting legs may be used to locate the transducers.

6.2.3 In cabins, vibration readings are to be taken in the centre of the floor area. For larger spaces, such as mess rooms and the navigating bridge, sufficient measurements should be taken to define the vibration characteristics.

6.3 Test conditions

6.3.1 Measurements should be taken with the vessel proceeding ahead at a constant speed and straight course in a depth of water not less than five times the draught.

6.3.2 Tests should be conducted in sea conditions not greater than sea state 3 (slight).

6.3.3 The vessel should be at a displacement and trim corresponding to a normal operating condition.

6.3.4 Rudder angle variations should be limited to ± 2° of the amidships position.

6.4 ISO 6954:1984 and ISO 6965:2000

6.4.1 ISO 6954:1984 Guidelines for the overall evaluation of vibration in merchant ships have been superseded by ISO 6954:2000. At the time of writing (2006) the latter version remains to be widely adopted. This reflects the considerable experience which exists among shipowners, shipbuilders, operators and Classification Societies with regard to the application and interpretation of the earlier version. This suggests that ISO 6954:1984 will remain in general use as a specification requirement until confidence has been established in ISO 6954:2000.

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6.4.2 Assessment criteria and limits for both standards are given below. The measurement parameters, frequency range, analysis methods, and severity criteria are different for the two versions and care must be taken to define which one is being used.

6.5 Lloyd’s Register’s assessment of accommodation and workspaces based on ISO

6954:1984

6.5.1 The guidance values are based on ISO 6954:1984 Mechanical Vibration and shock -Guidelines for

the overall evaluation of vibration in Merchant Ships (1984). The limits of this Standard have

been adopted as Figure 6.1 of the Guidance Notes.

The amendments and interpretation are intended to facilitate: simpler analysis of routine survey measurements; evaluation of the sum of two or more frequency components; and differentiate between rest and working areas.

6.5.2 The severity criteria are intended to apply to the position of maximum vibration within the space under consideration. Vibrations in each of the three orthogonal directions should be assessed independently. Analysis of measurements should be in accordance with

Section 3.2.

6.5.3 Broadband measurements.

The following severity classification applies to broadband measurements of the maximum overall vibration amplitude (peak) for vibrations within the frequency range 1 to 100 Hz. The values are the maximum repetitive amplitudes which occur during a minimum measuring period of 60 seconds.

Vibration levels are normal if:

• Velocity is below ± 6 mm/s peak,

or

• Acceleration is below ± 0.15 m/s2_peak.

Vibration levels are excessive if:

• Velocity exceeds ± 15 mm/s peak,

and

• Acceleration exceeds ± 0.30 m/s2_peak.

For intermediate vibration levels severity should be assessed from a narrowband frequency analysis as described in Paragraph 6.5.3.

6.5.4 Narrowband measurements.

Broadband measurements of vibration having levels which fall between the normal and excessive levels given in Paragraph 6.5.2 should be assessed using a narrowband frequency analysis in relation to the guidance values shown in Figure 6.1. Values are the maximum amplitudes of individual narrowband components which occur during a minimum measuring period of 60 seconds.

No single component should exceed the “upper boundary”. For sleeping and rest areas only one component may be within the stippled zone.

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Figure 6.1 Accommodation and workplace vibration levels.

Adverse comments not probable Adverse comments probable

1 1 mm/s 0. 01 m /s² 10 mm/s V e lo ci ty ( p e a k) , ± m m /s Frequency, Hz 10 0.1 m /s² 100 mm/s 0.1 m m 1 m /s_² 100 0.01 m m 10 m /s² Disp lacem ent, ± mm Acc eler at ion, ± m/ s² 100 m /s² 1 m m 10 m m Lloyd’s Register 30

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6.6 Assessment accommodation and workspaces based on ISO 6954:2000

6.6.1 Guidance values are based on ISO 6954:2000 Guidelines for the overall evaluation of vibration in merchant ships.

6.6.2 The severity criteria are intended to apply to the position of maximum vibration within the space under consideration. Vibrations in each of the three orthogonal directions should be assessed independently. Analysis of measurements should be in accordance with

Section 3.3.

6.6.3 The recommended maximum acceptable values of frequency weighted r.m.s. vibration velocity are given in Table 6.1. for vibration within the frequency range 1 to 80 Hz.

Table 6.1 Accommodation and workplace vibration levels.

Normal Excessive

Frequency weighted r.m.s. mm/s.

Crew accommodation areas 4.5 6

Working areas 6 8

6.7 Other ship types, PCAC

6.7.1 Vibration habitability criteria for passenger ships, high speed craft and yachts are given in LR’s Provisional Rules for Passenger and Crew Accommodation Comfort, (PCAC).

6.7.2 These Rules define the vibration severities in all relevant areas of the vessel for classifications of passenger comfort. Additionally, criteria are given for crew accommodation and working spaces which can be applied to conventional merchant ships where an enhanced comfort standard is required in relation to the basic acceptability limits given above in Sections 6.4, 6.5 and 6.6.

6.7.3 The PCAC Rules also stipulate more comprehensive test conditions and survey procedures that must be followed if a Class comfort notation is assigned to the vessel.

6.8 Motion sickness

Motion sickness is influenced by a number of factors such as translational and rotational oscillations, constant speed rotations about an off-vertical axis, Coriolis stimulation and movements of the visual scene. Whilst some types of motion can be reliably predicted to result in more nausea than others, motion sickness is neither explained nor predicted by dynamics. Biological factors such as age and sex also have an influence.

Acceleration, exposure time, and frequency of motion can be used to derive motion sickness indices, such as the Motion Sickness Dose Value (MSDV) described in ISO 2631. Motion sickness may be expected to occur in some unacclimatized people at accelerations greater than 2 to 3 m/s2_{at roll periods of around 10 seconds for normal ship type hulls.}

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Section 7. Machinery vibration

7.1 Introduction.

7.1.1 Measurements should be taken with the machine operating at normal temperatures.

7.1.2 The assessment criteria apply to all operating speeds and loads for which the running conditions are stable.

7.1.3 With the exception of main propulsion machinery, broadband measurements may be used for assessment if the vibration severity with the machinery operating is greater than three times the background value with the machine stopped. Otherwise narrowband analysis is required.

7.1.4 Narrowband severity should be assessed on the amplitudes of the frequency components generated by the machine itself.

7.2 Rotating machines

7.2.1 Assessment criteria

The assessment criteria are given in Table 7.1.

Table 7.1 Rotating machinery assessment

Satisfactory Unsatisfactory Small machines 1.0 0.2 20.0 10.0 0.71 1.12 Good Good <15 kW Medium machines 15-75 kW 1.8 4.5 2.8 7.1 Excessive Satisfactory Unsatisfactory Excessive 50.0 Satisfactory Unsatisfactory Excessive foundations Good Rigid foundations Flexible Large machines 1.8 4.5 11.2 2.8 7.1 18.0 Satisfactory Unsatisfactory Excessive Good

Class I Class II Class III Class IV

O ver al l vi br at ion v e lo ci ty , m m /s ec r. m .s. Lloyd’s Register 32

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7.2.2 Scope

Assessment of vibration severity for rotating machines above 10 kW.

7.2.3 Vibration severity.

The severity of rotating machinery vibration is defined by the root mean square (r.m.s.) value of velocity for frequencies up to 1000 Hz.

7.2.4 Position and direction

Measurements are to be taken at or adjacent to each main bearing in two orthogonal radial directions and in the axial direction, all referred to the shaft axis. Care must be taken that the measurements do not include local resonance effects.

The assessment categories given in Table 7.1 apply independently to each measuring direction.

7.2.5 Machine classes

Class I: Individual parts of machines, integrally connected with the complete machine in its normal operating condition. Production electric motors up to 15 kW are typical examples of machines in this category.

Class II: Medium sized machines (typically electrical motors with 15 to 75 kW output) without special foundations, rigidly mounted engines or machines (up to 300 kW) on special foundations.

Class III: Large prime movers and other large machines with rotating masses mounted on rigid and heavy foundations which are relatively stiff in the direction of vibration

measurement.

Class IV: Large prime movers and other large machines with rotating masses mounted on foundations which are relatively soft in the direction of vibration measurement (for example turbo-generator sets, especially those with light weight sub-structures).

7.2.6 Rigid and flexible supports

The distinction between “rigid” and “flexible” supports is defined as follows:

• Rigid support: the natural frequency of the predominant vibration response mode is greater than the excitation frequency

• Flexible support: the natural frequency of the predominant vibration response mode is less than the excitation frequency.

Where the support stiffness cannot be classified the assessment should be based on the “rigid support” values.

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7.2.7 Guidance on assessment categories

Good. Typical of new machines.

Satisfactory. Satisfactory for long-term operation.

Unsatisfactory. Unsatisfactory for long-term operation. The machine can be operated for limited periods until repairs are possible.

Excessive. Probable damage to the machine.

7.3 Reciprocating machines

7.3.1 Scope

Assessment of vibration severity on rigid and flexibly mounted reciprocating machines above 20 kW. The frequency range considered is 2 to 100 Hz.

7.3.2 Vibration severity

The severity of reciprocating machinery vibrations is defined by the peak vibration amplitudes.

7.3.3 Position and direction

Measurements are to be taken on top of the machine frame (entablature) in three mutually orthogonal directions referred to the shaft axis. The assessment criteria should be applied to the position and direction of the maximum vibration amplitude. The principal modes of vibration of slow speed diesel engines are shown in Figure 7.1.

Figure 7.1 Vibration modes of slow speed diesel engines.

X-mode H-mode Doub le bo ttom Doub le botto m Lloyd’s Register 34

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Figure 7.2 Reciprocating machinery.

Damage probable 100 mm/s 1 mm/s 0. 01 m /s² 1 Vel o c ity ( p eak ), ± m m /s 10 mm/s Frequency, Hz 10 0.1 m /s² Recommended 10 m m 1 m /s ² 100 10 m /s² 0.1 m m 0.01 m m 100 m_/s ² 1 m m Acc el_er at ion_{, ±} m/ s² Displ acem ent, ± mm Lloyd’s Register 35

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7.3.4 Assessment

Overall peak amplitudes for vibrations in the frequency range 2 to 100 Hz should be less than each of the following values:

• displacement ± 0.4 mm • velocity ± 25 mm/s • acceleration ± 40 m/s2_.

7.3.5 The assessment of higher vibration levels should be based on a narrowband frequency analysis referred to Figure 7.2. The assessment levels given in Figure 7.2 apply

independently to each measurement. Displacement limits below 15 Hz only apply to rigidly mounted machines.Resiliently mounted reciprocating machines.

The levels given in Figure 7.2 are applicable for the normal running condition. Large displacements are normal during starting and stopping when the machine speed coincides with the natural frequency of the resilient mounts. The Guidance Notes do not apply to this transitory condition.

7.4 Diesel alternator sets

7.4.1 The assessment criteria apply to both rigidly mounted and resiliently mounted sets.

7.4.2 Vibration levels on alternators driven by reciprocating internal combustion engine should be assessed using Table 7.2 which have been reproduced from ISO 8528 Part 9 table C.1.

7.4.3 Alternator Assessment.

• <value 1. Damage not expected with standard designs if vibration less than this level.

• >value 1, <value 2. Assessment of structure and components may be required, along with an agreement between manufacturer and component supplier to ensure reliable operation.

• >value 2. Special designs of structure and components required.

7.4.4 Engine assessment.

In general, engine vibration levels which lie within the “Recommended” area of Figure 7.2 will also comply with the Value 1 of ISO 8528 Part 9.

7.5 Turbochargers

In the low frequency range (less than approximately 100 Hz) vibration levels on the turbocharger body should be assessed using Figure 7.2.

Vibration at higher frequencies relating to turbocharger rotor speed and orders thereof should be assessed using Table 7.1. Class II is suitable for small frame turbochargers. Class III is suitable for large frame turbochargers found on slow speed diesel engines.

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Table 7.2 Vibration of AC generating sets driven by reciprocating internal combustion engines.

Speed Power Displacement, r.m.s. Velocity, r.m.s. Acceleration, r.m.s.

Engine Generator Engine Generator Engine Generator

value 1 value 2 value 1 value 2 value 1 value 2 (rpm) (kW) 1) _{mm mm mm mm/s}_mm/s_mm/s_m/s2 _m/s2 _m/s2 ≤ 3600 > 40 0.64 2) _0.82) ₄₀2) ₅₀2) ₂₅2) ₃₁2) ≥ 2000 ≤ 40 0.8 0.95 50 60 31 38 < 2000 > 200 0.72 0.32 0.45 45 20 28 28 13 18 ≥ 1300 ≤ 200 > 100 0.72 0.4 0.48 45 25 30 28 16 19 ≤ 100 > 40 0.4 0.48 25 30 16 19 < 1300 > 720 > 1000 0.29 0.35 18 22 11 14 ≤ 1000 > 200 0.72 0.32 0.39 45 20 24 28 13 15 ≤ 720 > 1000 0.72 0.24 0.32 45 15 20 28 9.5 13 Notes:

1) Equivalent electrical power, kVA = cos ϕ .kW where cos ϕ = 0.8 (typically). 2) These values are subject to agreement between the manufacturer and customer.

3) Where generators are coupled to the flange housing of the engine the values measured at the engine end of the generator shall meet the values for generators.

7.6 Reciprocating compressors

The values given in Figure 7.2 for reciprocating machinery are applicable.

7.7 Main propulsion gearboxes

Main propulsion gearboxes should be assessed as rotating machinery using Table 7.1.

7.8 Rolling element bearings

Rolling element vibration measurements may require special instrumentation because the defect frequencies may be above the frequency response range of general purpose vibration instruments. Various parameters have been developed such as shock pulse, spike energy, curtosis factor, and acceleration crest factor. The measured values are dependent on the instrumentation used and there are no internationally accepted standards. The most useful technique for fault finding is considered to be the envelope analysis described in

Section 3.1.7. Otherwise, the recommended procedure is to make regular, periodic measurements and observe the trend of the levels.

7.9 Pipework

Pipe vibrations can be assessed in the first instance using Figure 7.3. The guidance values are based on experience ashore in the petrochemical industry.